The present invention relates to blades form turbomachinery and, more particularly, to an improved mechanism for dampening vibrations experienced by the blades or disks of turbomachinery so as to extend their operational lives and improve machine performance.
Turbomachinery is used in many applications to perform work on or extract work from both gaseous and liquid fluids. Examples of such machinery include gas turbines, axial and centrifugal fans, marine and aviation propellers, fan blades, helicopter blades, tail rotors, wind turbines, turbo pumps, and steam and hydraulic power turbines. This machinery may contain one or more broad class of rotating and fixed appendages including blades, vanes, foils, and impellers depending on the need of a particular machine.
Turbines and compressor sections within an axial flow turbine engine, as well as other turbomachinery, generally include a rotator assembly comprising a rotating disk and a plurality of rotor blades circumferentially disposed around the disk. During operation, turbine engine rotator assemblies rotate at a variety of speeds through fluid that varies in temperature, pressure, and density. As a result, the blades may be excited in a number of different modes of vibration. Lower order modes manifest themselves in bending modes and torsion modes, whereas higher order modes have more complex blade deformations.
The rotating blades are subjected to resonating conditions as more fully described in U.S. Pat. No. 5,924,845 (""845) which is herein incorporated by reference. As described in the ""845 patent, potentially destructive resonant vibration can occur when the frequency of an alternating excitation force imposed on a blade is near a natural, or resonant, frequency of the blade. At these resonant frequencies a blade will experience peak vibrations. A typical blade will have more than one resonant frequency, representing the various modes, or ways, in which the blade can vibrate. For example, a jet engine blade may be viewed as a simple cantilever beam. If an alternating excitation force is imposed on the blade at a resonant frequency corresponding to the first mode, it will simply bend back and forth with one wave along its length. If the excitation frequency is at the second mode resonant frequency, the blade will bend with two waves along its length, and so forth. Modern jet engine blades are more like plates than beams, so that the blades have more complicated vibration modes, including, in addition to conventional bending modes, torsion modes and chordwise bending modes. All of these vibration modes combine to determine the actual resonant frequencies for a turbomachine blade.
Excitation forces in turbomachines arise from time dependent fluid flows. For example, one usual source of an excitation force is the aerodynamic force imposed on a rotor blade each time it rotates past a stator vane (a stationary blade). There will be a number of stator vanes past which the rotor blade will travel during one rotor revolution, causing the frequency of aerodynamic excitation to be equal to the number of stator vanes multiplied by the rotor speed. This is called engine order excitation. Since there are generally several rows of stator vanes and rotor blades of varying numbers, there are typically many engine order excitation frequencies that a rotor blade will experience. Each of these physical sources of vibratory excitation will cause a different so-called speed line, which is an integer multiple of engine rotor speed.
As first described by W. Campbell in a pioneering 1924 work describing the problems of vibration in turbomachinery, this can be described in a Campbell diagram, where speed lines are plotted on a graph as functions of the rotor speed. Also plotted on the Campbell diagram are the various frequencies at which resonant vibration will occur for each mode of the rotor blade in question. Wherever a speed line, corresponding to a regularly occurring vibratory excitation in a turbomachine, crosses a resonant blade frequency line, resonant blade vibration is possible. This speed line can cross several blade resonant frequencies as the turbomachine speeds up. This means that the excitation frequency on blades will coincide with successively higher resonant frequencies of the blades as rotor speed increases.
Adding one or more dampening devices to a rotor blade will decrease the vibrations and lessen the damage to the blade. In addition, it may allow the turbomachine to be run at a higher speed, improving the efficiency of the machine. Dampening devices may also be added to the rotor disk, which may have vibration modes coupled to the rotor blades. The dampers can then reduce vibrations in the rotor disk and/or the rotor blade.
Rotor blades that are used in axial flow turbines and compressors can be excited to severe levels of vibration when subjected to time-dependent forces as described above. The severe vibrations can damage the material in rotor blades. Devices that absorb and dissipate vibration energy in the blades of rotors in compressors and/or turbines, such as the centrifugal pendulum absorber disclosed in the ""845 patent, are known and more of which are disclosed in U.S. Pat. Nos. 4,182,599; 4,360,088; 4,441,859; 4,484,859; 4,650,167; 5,052,890; 5,232,344; 5,346,362; 5,369,882; 5,498,137; 5,749,705; and 5,820,348 all of which are herein incorporated by reference. It is desired that further improvements for a vibration dampening mechanism for blades be provided
It is the primary object of the present invention to provide for a dampening mechanism that absorbs and dissipates vibration energy in the blades or disks of compressors and/or turbines so to extend the service life thereof.
It is another object of the present invention to provide for a dampening mechanism that uses the technique of self-tuned impact dampening.
It is another object of the present invention to provide for a self-tuned impact dampening mechanism that is adjustable to dampen engine-order resonant vibrations along a speed line.
It is a further object of the present invention to provide for a self-tuning dampening mechanism that utilizes a ball-in-spherical trough configuration.
It is a still further object of the present invention to provide for a self-tuning dampening mechanism that can be located internal or external to the blade for which it provides the dampening features so as to extend its operational life.
This invention is directed to a self-tuning dampening mechanism for a rotating appendage for a turbomachine which absorbs and dissipates the vibration energy in the appendage so as to extend its operational life.
The vibration damper device comprises a member coupled to a rotor having a frequency of vibration with the member being subjected to fluid-flow forces when the rotor is rotated. The member has one or more cavities with walls which confine a rattling mass in each cavity having parameters that are selected in accordance with the anticipated frequency of aerodynamic excitation associated with a speed line on the appropriate Campbell diagram.
In one embodiment, the vibration damper device has a trough configuration having a spherical bottom and a rattling mass, such as a bail, having a resonant frequency that corresponds to the anticipated frequency of aerodynamic excitation.
For a better understanding of the nature and objects of the present invention, reference should be made to the following detailed description taken in conjunction with the accompanied drawings, in which like parts are given like reference numbers, and wherein:
FIG. 1 is a schematic drawing of a turbine wheel with blades and a shaft;
FIG. 2 is composed of FIGS. 2(A), 2(B), and 2(C), wherein FIGS. 2(A), and 2(B) schematically illustrate prior art dampening mechanisms, and FIG. 2(C) schematically illustrates the self-tuning impact damper mechanism of the present invention;
FIG. 3 illustrates a Campbell diagram showing speed lines and resonant frequencies plotted against rotor speed;
FIG. 4 illustrates a two (2) degree-of-freedom system;
FIG. 5 illustrates various curves associated with tuned mass amplification factor calculations;
FIG. 6 is composed of FIGS. 6(A) and 6(B) that respectively illustrate the amplification factor calculation curves and damping factor calculation curves associated with the self-tuning impact damper of the present invention;
FIG. 7 illustrates response curves associated with the tuned mass damper effectiveness;
FIG. 8 illustrates response curves associated with the self-tuning impact damper effectiveness of the present invention;
FIG. 9 is composed of FIGS. 9(A) and 9(B) that illustrate various views of the ball-in-spherical trough configuration of the present invention;
FIG. 10 illustrates the test plates used during the performance of testing related to the present invention;
FIG. 11 illustrates the experimental results obtained from the self-tuning impact damper mechanism of the present invention; and
FIG. 12 illustrates further experimental results obtained from the self-tuning impact damper mechanism of the present invention.